Methods for fluid separations, and devices capable of separating fluids

ABSTRACT

Wick-Containing apparatus capable of separating fluids and methods of separating fluids using wicks are disclosed.

RELATED APPLICATIONS

This application is a divisional of U.S. patent application Ser. No.10/422,367, filed Apr. 23, 2003 that now U.S. Pat. No. 7,051,540, inaccordance with 35 U.S.C. sect. 119(e), claimed priority U.S.Provisional Application Nos. 60/443,070, filed January 2003, and60/451,880, filed Mar. 3, 2003.

This invention was made with Government support under ContractDE-AC0676RLO1830 awarded by the U.S. Department of Energy. TheGovernment has certain rights in the invention.

FIELD OF THE INVENTION

The present invention relates to methods of separating fluids. Theinvention also relates to devices that are capable of separating fluids.

INTRODUCTION

Condensation and phase separation are important unit operations in manyprocesses, including space applications, such as water management inenvironmental life support systems and space suits (Lange and Lin,1998). As another example, fuel processors for fuel cells that convert aliquid hydrocarbon to hydrogen rich gas for fuel cells, are heavyconsumers of water (Flynn, et al., 1999), although net water is producedwhen coupled to a fuel cell. The ability to recover and recycle water iscritical in fuel cell systems to reduce the mass of consumables,particularly in transportation and portable applications. For both ofthese applications, size and weight of the hardware are criticalconsiderations. Furthermore, the ability to cool with a gas rather thana liquid is a significant advantage. A third application for compactcondensers is in portable cooling systems for soldiers and emergencyworkers.

Channels having a minimum dimension between 100 microns and a fewmillimeters can be used to accomplish phase separation in compactdevices (Wegeng, et al., 2001). Furthermore, hydrodynamic and capillaryforces have been shown to dominate over gravitational forces(TeGrotenhuis and Stenkamp, 2001), making these devices operableindependent of gravity and of orientation. Several other technologieshave been developed for phase separation in the absence of gravity, manyof which are rotary or vortex devices (Dean, 1991).

The development of compact heat exchangers is a well-established field(Kays and London, 1984, Webb, 1994), generating many techniques forenhancing heat exchange by reducing hydraulic diameter, adding extendedsurfaces, and inducing mixing.

SUMMARY OF THE INVENTION

In a first aspect, the invention provides a condenser comprising thefollowing elements in the order listed: a first cooling channel; a firstgas flow channel adjacent to the first cooling channel; a liquid flowpath comprising a wick; a second gas flow channel; and a second coolingchannel adjacent to the second gas flow channel.

The invention also provides a process of separating fluids in which afluid mixture passes into the first gas flow channel of the condenser ofthe first aspect.

In another aspect, the invention provides a condenser, that includes: acooling channel; a gas flow channel adjacent to the cooling channel; anda liquid flow path comprising a wick; wherein the liquid flow path isadjacent to the gas flow channel. The cooling channel is defined bycooling channel walls. This condenser possesses high energy densitysteady-state performance such that, when ambient air at 20° C. is passedthrough the cooling channel at a superficial velocity of 840 cm/s and afeed stream containing 40.0 mol % water vapor in air is passed throughthe gas flow channel at a superficial velocity of 1700 cm/s at theentrance, the decrease in pressure of the ambient air stream through thecooling channel is no more than 4 inches (10 cm) of water column, and atleast one of the following is obtained: (1) the energy density, ascalculated from the volume of the sum of the cooling channel and the gasflow channel, including the volume of walls defining the coolingchannel, is at least 2.0 W/cm³, or (2) the specific energy, ascalculated from the weight of the materials defining the cooling channeland the gas flow channel, is at least 1000 W/kg, or (3) the overall meanheat transfer coefficient is at least 500 W/cm²·K based on the primaryheat transfer area between the gas flow channel and the cooling channel,or (4) at least 70% of the water vapor in the feed stream condenses intoa liquid. The properties that characterize the condenser are to bemeasured at steady-state.

In a further aspect, the invention provides a method of condensingwater, comprising: passing a fluid mixture comprising water vapor into agas flow channel in a condenser, forming a liquid in a liquid flow path;and passing ambient air through a cooling channel with a pressure dropthrough the cooling channel of no more than 4 inches (10 cm) of watercolumn. The condenser comprises: a cooling channel defined by coolingchannel walls; a gas flow channel adjacent to the cooling channel; and aliquid flow path that is adjacent to the gas flow channel. There is aprimary heat transfer surface between the gas flow channel and thecooling channel. This method is further characterized by at least one ofthe following: (1) the energy density, calculated from the volume of thesum of the cooling channel and the gas flow channel, including thevolume of walls defining the cooling channel, is at least 2.0 W/cm³, or(2) the specific energy, calculated from the weight of the materialsdefining the cooling channel and the gas flow channel, is at least 1000W/kg, or (3) the overall mean heat transfer coefficient is at least 500W/cm²·K based on the primary heat transfer area between the gas flowchannel and the cooling channel, or (4) at least 70% of the water vaporin the feed stream condenses into a liquid.

Properties and experimental results are disclosed in the description ofpreferred embodiments and Examples sections. Processes and apparatus ofthe invention can, alternatively or additionally, be characterized bythe properties and results described. For example, methods of thepresent invention can be described in terms of pore throat utilization,heat flux, heat transfer coefficient, and energy density or specificpower as a function of water vapor in the feed stream. Similarly, theinventive apparatus can be described as being characterizable by valuesof pore throat utilization, heat flux, heat transfer coefficient, andenergy density or specific power as a function of water vapor in thefeed stream, if tested according to the conditions described in theExamples.

The invention also provides systems that incorporate any of thecondensers described herein. For example, the invention includes systemsin which a condenser is connected to the outlet of a fuel cell,combustor (preferably a microchannel combustor), or fuel processor (suchas a steam reforming unit). Likewise, some methods of the presentinvention perform the functions of these systems; for example, recoveryof liquid water from the effluent of a fuel cell, or recovery of waterfrom a combustion reaction, or recovery of water from a fuel processor.

In some preferred embodiments, any of the condensers, methods, andsystems described herein are gravity independent.

The invention, in various aspects and embodiments can provide numerousadvantages including: rapid mass transport, high rates of heat transfer,low cost, durability, highly efficient gas-liquid and fluid separationsin a compact space, low profile equipment, and unit process operationsthat function in the absence of gravity, such as in extraterrestrialapplications. The invention can also be advantageous in applicationswhere flow rates are small or where size matters, examples include:analytical systems, biological applications, waste stream purification,recovery and recycling such as urine in space applications.

GLOSSARY OF TERMS

“Breakthrough pressure” is the maximum pressure difference that can bemaintained across a porous structure without having a wetting fluiddisplaced from the porous structure by a second fluid.

A “capture structure” is a structure disposed (at least partly) within agas flow channel that assists movement of a liquid into the wick.

“Device volume” refers to the entire volume of the device, includingchannels, headers, and shims.

“Flow microchannel” refers to a microchannel through which a fluid flowsduring normal operation of an apparatus.

A “fluid mixture” comprises at least two components, one of which will(at least partially) form a liquid phase in a liquid flow path.Typically, a fluid mixture contains a condensable component (such asgaseous water) and a noncondensable component (such as N₂); however, afluid mixture could also be comprised of a gas (such as N₂) andsuspended liquid droplets (such as water droplets).

A “gas flow channel” may or may not contain a capture structure. In anycase, a gas flow channel contains less wicking material than in anadjacent liquid flow path so that a liquid will preferentially migrateto the liquid flow path.

A device, or method, that is “gravity independent” or “orientationindependent” functions well in the absence of gravity or in anyorientation with respect to a gravitational field. Motivating liquidflow via capillary forces can enable this mode of operation.

A “laminated device” is a device having at least two nonidenticallayers, wherein these at least two nonidentical layers can perform aunit operation, such as heat transfer, condensation, etc., and whereeach of the two nonidentical layers are capable having a fluid flowthrough the layer. In the present invention, a laminated device is not abundle of fibers in a fluid medium.

A “liquid” is a substance that is in the liquid phase within the wickunder the relevant operating conditions.

A “liquid flow path” is a wick (or wicks) or open channel (or channels)or pore throat (or pore throats) or a combination of wicks, porethroats, and open channels through which a liquid flows during operationof a device.

“Microchannel” refers to a channel having at least one dimension of 5 mmor less. The length of a microchannel is defined as the furthestdirection a fluid could flow, during normal operation, before hitting awall. The width and depth are perpendicular to length, and to eachother, and, in the illustrated embodiments, width is measured in theplane of a shim or layer.

“Microcomponent” is a component that, during operation, is part of aunit process operation and has a dimension that is 1 mm or less.

“Pore throat” refers to a porous structure having a maximum poredimension such that a non-wetting fluid is restricted from displacing awetting fluid contained with the pore throat under normal operatingconditions.

“Residence time” refers to the time that a fluid occupies a givenworking volume.

“Superficial velocity” is calculated as the volumetric flow rate of afluid divided by the total flow area available to the flowing fluid.

“Unit process operation” refers to an operation in which the chemical orphysical properties of a fluid stream are modified. Unit processoperations (also called unit operations) may include modifications in afluid stream's temperature, pressure or composition.

A “wicking region” is the volume occupied by a wick, or, a wickingsurface such as a grooved microchannel surface.

“Working volume” refers to the total channel volume of the device, andexcludes the headers and solid shim and end plate materials.

BRIEF DESCRIPTION OF THE FIGURES

FIG. 1 is a Schematic illustration of an exploded, cross-sectional viewof a phase separator.

FIG. 2 is a compilation of results at varying condensing stream air flowrates and pore throat utilization in the device of Example 1 whereliquid breakthrough occurred (o) and did not occur (x).

FIG. 3 is a compilation of results at varying condensing stream air flowrates and gas outlet temperatures in the device of Example 1 whereliquid breakthrough occurred (o) and did not occur (x).

FIG. 4 is a plot of average heat flux versus percent water in the feedat condensing stream air flows of 11 SLPM (+), 10 SLPM (⋄), 9 SLPM (Δ),8 SLPM (x), 7 SLPM (o) and 5 SLPM (▪), in the device of Example 1.

FIG. 5 is a plot of mean overall heat transfer coefficient versuspercent water in the feed at condensing stream air flows of 11 SLPM (+),10 SLPM (⋄), 9 SLPM (Δ), 8 SLPM (x), 7 SLPM (o) and 5 SLPM (▪), in thedevice of Example 1.

FIG. 6 is a plot of thermal energy density and specific power versuspercent water in the feed at condensing stream air flows of 11 SLPM (+),10 SLPM (⋄), 9 SLPM (Δ), 8 SLPM (x), 7 SLPM (o) and 5 SLPM (▪), in thedevice of Example 1.

FIG. 7 is a plot of percent water recovered versus the water content offluid entering the fluid inlet of the device of Example 2 at condensingstream air flows of 32 SLPM (●) and 48 SLPM (▪) with trendlines shownthrough the predicted performance values at the same operatingconditions that produced the experimental data for both 32 SLPM air flow(- - -) and 48 SLPM air flow (—).

FIG. 8 is a plot of specific power versus the water content of fluidentering the fluid inlet of the device of Example 2 at condensing streamair flows of 32 SLPM (●) and 48 SLPM (▪) with trendlines shown throughthe predicted performance values at the same operating conditions thatproduced the experimental data for both 32 SLPM air flow (- - -) and 48SLPM air flow (—).

FIG. 9 is a plot of average heat flux versus the water content of fluidentering the fluid inlet of the device of Example 2 at condensing streamair flows of 32 SLPM (●) and 48 SLPM (▪) with trendlines shown throughthe predicted performance values at the same operating conditions thatproduced the experimental data for both 32 SLPM air flow (- - -) and 48SLPM air flow (—).

DESCRIPTION OF PREFERRED EMBODIMENTS OF THE INVENTION

General features of laminated, capillary-driven fluid separation devicesand prefer-red operating conditions are disclosed in U.S. PublishedPatent Application 20020144600 A1, which is incorporated herein byreference as if reproduced in full below.

A cross-sectional, schematic view of a condenser is shown in FIG. 1.During operation, a fluid mixture passes in through fluid inlet 2 intoheader 4 where it is distributed into gas flow channels 6 and 6′.Coolant passes through elongated coolant slots 8 in cooling channellayer 10. The material surrounding the coolant slots are the coolingchannel walls. As the fluid mixture passes through the gas flowchannels, heat from the fluid is removed through primary heat exchangesurface 13 (this surface also is an exterior surface of a coolingchannel wall) and a liquid condenses from the fluid mixture, flows intowick 11, through optional pore throat 12 and into liquid flow channel14. The figure is an exploded view and shows a separation between thewick and the pore throat; however, in typical operation the optionalpore throat should contact the wick. The device can work under theinfluence of gravity, but, more typically, suction is applied to pullliquid out through liquid outlet 16. In a device with multiple liquidflow channels, an optional footer (not shown) may carry flow frommultiple liquid flow channels. Gas from the gas flow channels may passthrough an optional gas footer and out through gas outlet 20.

While the device in FIG. 1 shows only one configuration, it should beappreciated that numerous variations come within the purview of theinvention; for example, there may be multiple gas outlets, there may bepore throat in the gas outlet, etc. In the illustrated device, thecoolant slots 8 extend to the gas outlet; however, in preferredembodiments, the coolant slots (or, more generally, the flow paths forcoolant) do not extend to the gas outlet, but end at a point 22 prior tothe outlet. This configuration can reduce condensation near the gasoutlet and thus reduce the risk of liquid breakthrough to the gasoutlet. In some preferred embodiments, the coolant flow path ends at adistance away from the gas outlet that is at least 10% of the length ofthe adjacent gas flow channel 6.

In preferred embodiments, the condenser is a laminated device made bystacking thin plates. Such a device may be plumbed similarly to thelaminated constructions described in U.S. Published Patent Application20020144600 A1.

In the devices of the present invention, the primary heat transfersurfaces are the walls between the heat exchangers and the gas flowchannels. Walls between channels in the heat exchanger can act as heatexchange fins, and thus provide extended heat transfer surface area.Walls within the heat exchanger can also provide structural support. Theoptimal aspect ratio for the heat exchanger channels and the thicknessof the walls within a heat exchanger that are between the heat exchangerchannels depend on the thermal conductivity of the material and theconvective heat transfer coefficient on the coolant side. In somepreferred embodiments, channels for fluid flow in the heat exchangerhave a wall thickness between slots of less than 20 μm, and preferably,a channel width of 15 to 50 μm, and a preferred ratio of channel height(a direction perpendicular to flow; in a laminated device, the stackingdirection) to channel width (the dimension perpendicular to height andflow) of at least 2, more preferably at least 4.

Coolant fluid flowing through the coolant channels can be a liquid (forexample, water) or a gas. In some embodiments, a fan or blower moves gasthrough the cooling channels. In some preferred applications of thepresent invention, it is desired to use a gas as the heat exchangefluid. In this case, the majority of the heat transfer resistance can bein the heat exchange channel. Furthermore, the application may besensitive to the pressure drop of the heat exchange fluid. One exampleis water recovery from multiple streams in an automotive fuel processorfor generating a hydrogen rich gas stream for use in a fuel cell.Ambient airflow provided by a blower or fan would be the most convenientheat exchange fluid. In this case, the pressure head provided by theblower or fan would be limited. For applications such as these, aconfiguration with extended heat transfer surface in the heat exchangechannels is preferred.

The presence of wicks and optional pore throats and optional capturestructures are common to multiple embodiments of the invention. A wickis a material that will preferentially retain a wetting fluid bycapillary forces and through which there are multiple continuouschannels through which liquids may travel by capillary flow. Thechannels can be regularly or irregularly shaped. Liquid will migratethrough a dry wick, while liquid in a liquid-containing wick can betransported by applying a pressure differential, such as suction, to apart or parts of the wick. The capillary pore size in the wick can beselected based on the contact angle of the liquid and the intendedpressure gradient in the device, and the surface tension of the liquid.Preferably, the pressure differential across the wick during operationshould be less than the breakthrough pressure—the point at which gaswill intrude into the wick displacing the liquid—this will exclude gasfrom the wick.

A liquid preferentially resides in the wick due to surface forces, i.e.wettability, and is held there by interracial tension. Flooding canresult from exceeding the flow capacity of the device for wetting phasethrough the wick; the flow capacity is determined by the fluidproperties, the pore structure of the wick, the cross-sectional area forflow, and the pressure drop in the wick in the direction of flow.

The wick can be made of different materials depending on the liquid thatis intended to be transported through the wick. The wick could be auniform material, a mixture of materials, a composite material, or agradient material. For example, the wick could be graded by pore size orwettability to help drain liquid in a desired direction. Examples ofwick materials suitable for use in the invention include: sinteredmetals, metal screens, metal foams, polymer fibers including cellulosicfibers, or other wetting, porous materials. The capillary pore sizes inthe wick materials are preferably in the range of 10 nm to 1 mm, morepreferably 100 nm to 0.1 mm, where these sizes are the largest porediameters in the cross-section of a wick observed by scanning electronmicroscopy (SEM). In some preferred embodiments, the wick is, orincludes, a microchannel structure. Liquid in the microchannels migratesby capillary flow. The microchannels can be of any length, preferablythe microchannels have a depth of 1 to 1000 micrometers (μm), morepreferably 10 to 500 μm. Preferably the microchannels have a width of 1to 1000 μm, more preferably 10 to 100 μm. In a preferred embodiment, themicrochannels are microgrooves, that is, microchannels having a constantor decreasing width from the top to the bottom of the groove. In anotherembodiment, the microchannels form the mouth to a larger diameter porefor liquid transport.

For wicking materials, the objective is to provide materials that havehigh permeability and small pore structure, in order to obtain high flowrates while also supporting a significant pressure drop down the wick(the maximum pressure drop decreases with increasing pore size). Fordevices where liquid phase mass transfer limits processing throughput,the thinness of the wick material is also critical for processintensification. Preferably the thickness of a wick is less than 500micrometers (μm), more preferably less than 100 μm, and in someembodiments between 50 and 150 μm.

In operation of a device with a wick, the wick should not be flooded,and it is preferably not dry. A wet or saturated wick will effectivelytransport liquid through capillary to a low pressure zone, such as lowpressure created by suction. A pore throat may be added to a liquidoutlet to prevent gas flow out of the liquid exit.

Punctured and punctured/expanded foils provide superior results whenused as wicks and/or capture structures in fluid separation apparatus.Particularly preferred foils are UltraThin MicroGrid Precision-ExpandedFoils, available from Delker Corporation. These materials are made in aflattened form and a 3-dimensional expanded form. Although similar toconventional wire mesh screens, these materials are made from a singlethin sheet by punching an array of hole while pulling the material. Inthe flattened form the holes are an array of diamonds. In the expandedform, the filaments are in a regular tetrahedral configuration.

Both the flat and expanded foils have been tested for wicking propertiesas single sheets, in multiple stacked sheets, and with or without asolid backing sheet. In general, the wicking properties are muchsuperior to other materials tested (except for Fresnel lenses, discussedbelow), including conventional woven screens. Qualitatively, the Delkerfoils appear to have an order of magnitude higher wicking rate than anyother regular or random porous structure tested. In addition, the Delkermaterials can be made in a wide variety of thickness as small as 0.0015inch (1.5 mil) and from a variety of metals, including copper, aluminum,and nickel.

Fresnel lenses are another preferred form of wick. Wicks havingmicrochannels having depths of less that 100 μm, preferably 50 to 100 μmpromote rapid mass transfer.

A wick can also be prepared by laser machining grooves into a ceramictape in the green state. Such wicks can be made, for example, withgrooves less than 50 microns deep with openings less than 100 micronswide. These grooves are expected to have a rectangular shape. Ceramicwicks have a high surface energy, are chemically inert, and have hightemperature stability. Another material option is intermetallics formedfrom two or more metals placed in intimate contact during a bondingprocess and which combine to form an alloy, compound, or metal solution.Preferred intermetallics will have properties very similar to theceramic materials. An advantage of engineered structures is fine controlof the length-scale for mass transfer in the liquid phase, which isdesirable for mass transfer limited applications, such as gas absorptionand distillation.

In preferred embodiments, a wicking/pore throat structure provides aflow path for a separated liquid phase. Two mechanisms are desirable foroptimal device operation, a wicking mechanism and a mechanism forexcluding gas. The wicking mechanism is accomplished by a porousstricture that is wetting for the liquid in order to cause preferentialsorption, while having high permeability to provide flow capacity to theoutlet. The second mechanism prevents intrusion by the gas stream intothe liquid outlet channel, and can be accomplished using a pore throat.The bubble point of the pore throat, as dictated by the maximum poresize, contact angle, and surface tension of the liquid, determines themaximum allowable pressure differential between the gas and liquidoutlets. The wick and the pore throat can be embodied in the samecomponent or structure if a suitable structure is available having ahigh enough wicking flow capacity and small enough pores to give a highenough bubble point. In devices having both a wick and a pore throat,the pore throat has a relatively greater resistance to fluid flow.

The wick is preferably not permitted to dry out during operation sincethis could result in gas escaping through the wick. One approach foravoiding dryout is to add a flow restrictor in capillary contact withthe wick structure, such as a porous structure with a smaller pore sizethan the wick structure and limiting the magnitude of the suctionpressure such that the non-wetting phase(s) cannot displace the wettingphase from the flow restrictor. This type of restrictor is also known asa pore throat. In preferred embodiments, a pore throat is providedbetween the wick and the liquid flow channel and/or at the liquidoutlet. In some embodiments, the wick can have a small pore diametersuch that serves to transport fluids from the gas channel and alsoprevents gas intrusion, thus serving the dual purpose of a wick and apore throat.

A pore throat has a bubble point that is greater than the maximumpressure difference across the pore throat during operation. Thisprecludes intrusion of gas into the pore throat due to capillary forces(surface tension, wettability, and contact angle dependent). The porethroat should seal the liquid exit, so there should be a seal around thepore throat or the pore throat should cover the exit in order to preventgas from bypassing the pore throat. The pore throat is preferably verythin to maximize liquid flow through the pore throat at a give pressuredrop across the pore throat. In some embodiments, the pore throat has apore size that is less than half that of the wick and a thickness of 50%or less than the wick's thickness; and in some of these embodiments thepore throat has a pore size that is 20% or less that of the wick.Preferably, the pore throat is in capillary contact with the wickingmaterial to prevent gas from being trapped between the wick and the porethroat and blocking the exit.

A capture structure can be inserted (at least partly) within the gasflow channel, and in liquid contact with the wick. The capture structureassists in removing (capturing) a liquid from the gas stream. Oneexample of a capture structure are cones that protrude from the wick;liquid can condense on the cones and migrate into the wick—an example ofthis capture structure is shown in U.S. Pat. No. 3,289,752, incorporatedherein by reference. Other capture structures include inverted cones, aliquid-nonwetting porous structure having a pore size gradient with poresizes getting larger toward the wick, a liquid-wetting porous structurehaving a pore size gradient with pore sizes getting smaller toward thewick and fibers such as found in commercial demisters or filter media.Mechanisms for capturing dispersed liquid particles include impingement(due to flow around obstructions), Brownian capture (long residence timein high surface area structure), gravity, centrifugal forces (highcurvature in flow), or incorporating fields, such as electrical or sonicfields, to induce aerosol particle motion relative to the flow field.

Capture structures can also be useful as a structural element. A wiremesh screen can be placed in the gas channel above a pore throat, sothat if a device is bolted together, the screen provides a force againstthe pore throat such that it seals against the rubber gasket on the backside and also creates a seal between the rubber gasket and the oppositewall. One means of assembling a multi-channel microchannel device is tocreate a sandwich of alternating layers with gaskets providing seals. Inthis situation, a capture structure in the gas channels could becompressed to generate forces through the entire stack, therebyproviding for the necessary seals.

Another use for a capture structure is to enhance heat transfer. If thecapture structure has a high thermal conductivity, it can act as anextended surface for heat transfer. This is advantageous where heattransfer is important, such as in condensation of at least some part ofa gas stream being cooled. By being in thermal contact with the primaryheat transfer area, the capture structure promotes heat removal from theflowing gas stream, which is then conducted to the primary area andsubsequently to the heat transfer fluid. In addition, condensation canoccur on the capture structure, and the heat of condensation can alsoconduct through the capture structure.

A further use for perforated foils in the expanded form (tetrahedrallyconfigured filaments) is as capture structures. Low flow resistance is adesirable attribute of a capture structure, and the open, regularstructure of Delker expanded screens (such as 10 AL 16-125 P and 5 Cu14-125 P) has low pressure drop for convective flow. The Delker foilscan have one to two orders of magnitude higher permeability thanconventional woven screens. In addition, the aluminum, copper, and othermetal forms have relatively high thermal conductivity and also enhanceheat transfer. Thus, tetrahedrally configured filaments provide asignificant advantage as capture structures.

A factor that may limit the throughput of the devices is the flowcapacity of the wicking and pore throat structure. This porous structureis characterized by a permeability coefficient defined as,

$\begin{matrix}{K = \frac{\mu_{L}{hQ}}{A\;\Delta\; P}} & (1)\end{matrix}$where Q is the volumetric flow of fluid through the cross-sectional areaA, of a porous media of thickness h, under an applied pressure drop ofΔP. The pore throat maximum liquid flow capacity, Q_(pt), is thencalculated for a given experiment from the viscosity of the liquid andthe pressure difference across the pore throat structure. Deviceperformance can then be characterized by the volumetric flow ofrecovered liquid as a percentage of the pore throat maximum flowcapacity. Preferably, in the present invention, the volumetric flow ofrecovered liquid as a percentage of the pore throat maximum flowcapacity is no more than 30%, more preferably no more than 10%.

Another potential limiting factor is intrusion of gas into the porethroat, which can occur at the bubble point of the pore throat, which iscalculated from the Young-Laplace equation,

$\begin{matrix}{{\Delta\; P_{m\;{ax}}} = \frac{2\;\sigma\;{\cos(\theta)}}{r_{p}}} & (2)\end{matrix}$where θ is the receding contact angle (meaning as the liquid is recedingacross the pore throat material) between the liquid and the pore throatand r_(p) is the maximum pore radius.

Since the wall separating the coolant flow channel from the gas flowchannel is the primary heat transfer surface, it may be advantageous toprovide a structure to aid the flow of liquid from the primary heattransfer surface to a wick in the liquid flow path. A capture structureor other transport structure can be used to provide a flow path to theliquid flow channel. A “transport structure” extends from the liquidflow path into the gas flow channel to either (1) the primary heattransfer surface, or (2) to a point near the primary heat transfersurface such that a liquid condensed on the primary heat transfersurface can flow along the structure into the liquid flow path.

In a device with an essentially planar gas flow channel, liquidcondensed on the primary heat transfer surface will be pushed to thesides of the gas flow channel under certain flow conditions (see thediscussion of Suratmann number in U.S. Published Patent Application20020144600 A1) where the liquid then comes in contact with and flowsinto a wick in the liquid flow path. In some preferred embodiments, thedistance between the primary heat transfer surface and a wick in aliquid flow path is 5 mm or less, more preferably 2 mm or less. In someembodiments, the areas for heat transfer and wick surface area arebalanced such that the area of the primary heat transfer surface iswithin 25% of the geometric surface area of the wick that is on theopposite side of the gas flow channel. In preferred embodiments, liquidbreakthrough out the end of a gas flow channel is reduced or preventedby limiting the area for cooling; thus, in some preferred embodiments,the coolant flow paths (the volume of a cooling channel layer thatcarries coolant) do not extend to the end of the gas flow channel, insome preferred embodiments, the coolant flow paths are not adjacent toat least 10% (in some embodiments 20%) of the length gas flow channelthat is nearest a gas outlet that is connected to the gas flow channel.

Another optional feature is reduced or non-wettability of the gas flowchannel wall adjacent to a heat exchange surface to preclude formationof a liquid film. This could be accomplished, for example, by makingthis wall of, or coating the wall with, a material that has a reduced ornon-wettability for the condensed phase (e.g., a hydrophobic materialwhere water is the condensed phase). The heat transfer coefficient wouldincrease substantially by avoiding the resistance of a liquid film.

The height of the gas flow channels 6, from surface 13 to wick surface11 is preferably about 10 μm to 5 mm, more preferably 100 μm to 1 mm.The height of the channels is preferably small for good heat and masstransfer and overall device size, balanced against potentially slowerflow rates or higher pressure drops. A high ratio of surface area ofexposed wick to volume of gas flow channel is desirable for efficientphase separations. Preferably this ratio is from 1 to 1000 cm²:cm³, andin some embodiments from 5 to 10. In some preferred embodiments, the gasflow channels are substantially the same length and the liquid flowchannel(s) is at least 10% shorter than the gas flow channels.

The illustrated embodiments show cross-flow heat exchange to provide forshorter coolant flow path and less coolant stream pressure drop;however, in some preferred embodiments, the flow through the heatexchanger is rotated 90° (so that the heat exchange fluid flows in thedirection opposite net fluid flow in the gas flow channels) to obtaincounter-current flow and higher heat transfer effectiveness. As in allthe devices described herein, the shims can be repeated for numerouslayers, and, in some embodiments, the devices include 2 to 1000, or atleast 4, repeating heat exchange units, where the repeating unitincludes shims for fluid separation (including a wick and capability forfluid transport) and heat exchange.

In some preferred embodiments there are multiple gas flow channelsoperating in parallel. This configuration allows high throughput andprovides a large surface area to volume ratio for high efficiency. Insome preferred embodiments, layers are stacked to have between 2 and 600separate gas flow channels, more preferably at least 3 gas flowchannels, and in some embodiments, between 3 and 40 gas flow channels.As an alternative to the parallel arrangement, the channels could beconnected in series to create a longer flow path.

In some preferred embodiments, net flow of coolant through the coolantflow path is cross-flow with respect to the net flow of gas through thegas flow channel; in this configuration, the length of the gas flowchannel (in the direction of net flow of gas) is preferably at least 2times, in some embodiments at least 5 times, longer than the length (inthe direction of net flow of coolant) of the coolant channels. Thisconfiguration can achieve a low pressure drop. Regardless of whetherthere is cross-flow, in some preferred embodiments, pressure dropthrough a coolant channel (meaning from the beginning to end of acoolant channel) is 4 inches (10 cm) of water column or less, and insome embodiments, 2 inches (5 cm) of water column or less.

Another advantage of some preferred embodiments of the invention is thatthe gas flow channels and/or liquid flow channels can be essentiallyplanar in the fluid separation regions. This configuration enableshighly rapid and uniform rates of mass and heat transport. In somepreferred embodiments, the gas flow channels and/or liquid flow channelshave dimensions of width and length that are at least 10 times largerthan the dimension of height (which is perpendicular to net gas flow).In especially preferred embodiments, the devices are made by stackingplanar shims (plates) and bonding the stacked shims. Preferably, theshims are less than 1 cm thick, more preferably less than 5 mm thick.

The effectiveness at preventing breakthrough of liquid into the gasoutlet is sensitive to two dimensionless parameters, the ratio of thegas and liquid Reynolds number and the Suratmann number. The Reynoldsnumber for both phases is calculated based on the space velocity for thesingle phase. The Suratmann number is defined as Su=σD_(h)ρ_(L)/μ_(L)with σ being the gas-liquid interfacial tension, D_(h) being thehydraulic diameter, ρ_(L) the liquid phase density, and μ_(L) the liquidphase viscosity. These two parameters have been identified in theliterature as indicating where the transition from annular to slug flowoccurs in pipe flow in microgravity, see Jayawardena, S., V.Balakotaiah, and L. C. Witte, “Flow Pattern Transition Maps forMicrogravity Two-Phase Flows”, AIChE J., 43(6), 1637-1640, 1997. Thedevice would operate better in the annular flow regime than in a slugflow regime, because in annular flow the liquid would be forced to thecorners and walls where it could then drain from the gas channel intothe wick structure. Conditions in the gas flow channel(s) are preferablymaintained such that Re_(GS)/Re_(LS) is greater than about(4500)·(Su)^(−0.67); and in some embodiments, the range ofRe_(GS)/Re_(LS) is in the range of (4600 to 10,0000)·(Su)^(−0.67).

In a preferred embodiment, the gas phase is contiguously connected tothe gas outlet and the liquid phase is contiguously connected from theliquid flow path to the liquid outlet. The continuity of phases at thegas outlet is effected by the geometry, the total flow and ratio of gasto liquid flow, and the fluid physical properties, as reflected by thedependence on Re_(GS)/Re_(Ls) and the Suratmann number described above.A second desired condition is sufficient wicking capacity, which isinfluenced by the flow area, fluid physical properties, and thepermeability of the material.

An inverse relationship has been discovered between the requirement forexcess flow capacity in the wick and the establishment of continuousphases in the gas flow channel. For a given gas and liquid flow rate,the size of the channel and the number of channels can be designed toachieve continuous phases (annular or stratified, as examples) in theflow in the channels as determined by the flow rates, the geometry, andthe physical properties of the fluids, including consideration of theratio of the Reynolds number and the value of the Suratmann number,while maximizing the throughput of the device.

In some preferred embodiments, the inventive device is characterized byany of the measurements described in the following Examples section. Forexample, in some preferred embodiments, the device possesses high energydensity performance such that, when air at 20° C. is passed throughcooling channels at a superficial velocity of 2100 cm/s and a feedcontaining 40.0 mol % water vapor in air enters gas flow channels at asuperficial velocity of 630 cm/s, the energy density calculated from thevolume of the sum of the volume of the cooling channel and the gas flowchannel, including the volume of walls defining the cooling channel, isat least 1 W/cm³, more preferably at least 4 W/cm³, and in someembodiments between about 1 W/cm³ and about 5 W/cm³. Using the weight ofmaterial (of the same volume described above for energy density), thespecific heat transfer power density is at least 1000 W/kg, morepreferably at least 3000 W/kg, and in some embodiments between about1000 W/kg and about 4000 W/kg. The heat flux and calculated overall meanheat transfer coefficient based on the primary heat transfer area (thearea of the wall separating coolant flow path from gas flow channel) areat least 1 W/cm² and 500 W/cm²·K, respectively, and more preferably atleast 3 W/cm² and 800 W/cm²·K, respectively. In some embodiments, theheat flux would be between about 1 W/cm² and 5 W/cm², with overall meanheat transfer coefficients between about 500 W/cm²·K and 1000 W/cm²·K.Furthermore, the decrease in pressure of the coolant stream would bepreferably no more than 10 inches (25 cm) of water column, morepreferably less than 4 inches (10 cm) of water column, and in someembodiments between 1 inch (2.5 cm) of water column and 5 inches (13 cm)of water column. In devices with multiple repeating units, theseproperties may be averaged over some or all of the repeating units. Insome inventive embodiments, the temperature of the coolant entering thecooling channels is at least 15° C. The pressure decrease of the coolantstream should be measured from the beginning of the coolant flow path orpaths that is adjacent to a gas flow channel or channels to the end ofthe coolant flow path or paths that is adjacent to a gas flow channel orchannels. Similarly, the device can be characterized by any of themeasured values (or at least 30% of the measured values, or from about40% of the measured value to about 100% of the measured value). In viewof the parameters described in these descriptions, skilled engineers can(through routine experimentation) optimize performance over thatdescribed in the Examples section.

EXAMPLES Example 1

The microchannel condenser is a cross-flow, air-cooled heat exchangerwith an integrated phase separator to collect and remove the condensateas a separate liquid stream from the device.

The device is an assembly of stacked components. An exploded-viewschematic of the stack is shown in FIG. 1. On the bottom is a heatexchange element containing slots for cooling air flow. The separatorelement is placed on top of the heat exchange element with a gasket inbetween thereby forming flow channels for the condensing stream. Asecond gasket and condensing element is placed on top of the separatorforming a second set of condensing channels. A mixture of vapor andnon-condensable gas is fed into a header region at one end anddistributed into an array of microchannels formed by the heat exchangeelement on one side and a wicking structure on the other. Air blowingcross-current through vertical slots in the heat exchange elements,cools the gas mixture and condenses the vapor flowing through themicrochannels. The objective is to sorb the condensate into the adjacentwicking structure and preclude the build-up of a liquid film on thecondensing surface that would add to heat transfer resistance. A premiseof the design is that by operating the device in the appropriate flowregime (TeGrotenhuis and Stenkamp, 2001), the liquid can be effectivelytransported from the condensing surface to the wick without entrainmentof liquid in the flowing gas leading to breakthrough of liquid to thegas outlet.

The phase separator located between the condensing microchannels iscomprised of two wicking structures and a liquid flow channel.Condensate flows through the wicks to pore throat windows that permitliquid flow but preclude gas flow into a liquid collection channel. Atube connected to one end of the liquid channel allows condensate to beremoved from the device.

The stack is compressed slightly within a housing to seal the condensingflow channels. The housing is connected to a blower to supply coolingair flow. Separate connections are also provided for hot stream feed,gas outlet flow, and a liquid condensate outlet. The heat exchangeaspects are described first followed by a description of the integratedphase separator.

The heat exchange elements are aluminum and weigh 23.5 g. Cooling isprovided by air flow through 142 slots that are 0.024 inch wide (0.61mm) by 0.100 inch tall (2.5 mm) and 1.2 inches long (3.05 cm). The slotsare separated by 0.010 inch wide (0.25 mm) walls that serve as heatexchange fins to enhance cooling side heat exchange. Calculated fineffectiveness is 99%. The cross-flow device is designed to operate withless than 10 inches of water column (2.5 kPa) pressure drop in thecooling air stream to facilitate the use of a fan or blower.

The flow channels for the condensing side of the heat exchanger aremicrochannels formed by the ridges located on top of the condensingelement and the separator element stacked on top of the condensingelement. The channels are 0.020 inches deep (0.5 mm) and 0.170 inch wide(4.3 mm) and 5.7 inches long (14.5 cm). There are five channels formedby each condensing element for a total of ten microchannels forcondensing heat exchange. The ridges separating the condensing channelsare structural and do not significantly enhance hot side heat exchange.The hot stream is introduced through a half-pipe tee into the inletheader at one end of the device and the uncondensed gas is collected inthe header at the opposite and exits through a second half-pipe tee.

The wall separating the hot side from the cold side provides the primaryheat exchange area and is 0.020 inch thick (0.5 mm). The total primaryheat exchange area on the condensing side is 62.2 cm². The primary areaon the cooling side of 44 cm² is enhanced to 410 cm² by the fins.

The phase separator is located between the two sets of condensing flowchannels. A liquid flow channel is isolated from the condensing flowchannels by pore throat material made of sintered stainless steel. Theseparator has three pore throat windows that are 1 inch by 0.5 inches(2.5 cm by 1.3 cm) on each side facing the condensing channels. The porethroat material is wetting for the condensate phase and allowscondensate flow from the condensing channels to the liquid flow channel,which occurs by maintaining the liquid outlet at a lower pressure thanthe gas outlet. Gas intrusion through the pore throat into the liquidflow channel is precluded by capillary forces, thereby promoting phaseseparation. The pore throat is characterized by permeability andbreakthrough pressure, the pressure difference where gas will displaceliquid from the pore throat. The average measured permeability of the0.0279 μm thick pore throat material in the separator is 6×10 ⁻¹⁰ cm²,and the breakthrough pressure is approximately 17 inches of water column(4.4 kPa).

Phase separation is further augmented by placing sorbent material (acotton cloth purchased from a fabric store) over the surface of thephase separator adjacent to the condensing flow channels. The purpose ofthe sorbent material is to locally collect condensate from the flowchannels and conduct the condensate to the pore throat windows where itflows through to the liquid flow channel and subsequently from thedevice.

Performance of the partial condenser and phase separator was tested onNASA's KC-135 reduced gravity aircraft. Testing occurred as the aircraftperforms parabolic maneuvers every 1 to 1.5 minutes thereby achievingshort periods of reduced gravity, typically at less than 0.04 g, lastingabout 20 seconds. Parabolas were performed in sets of about 10 separatedby one to three minute breaks as the aircraft turned around. Typically,40 parabolas were performed during a given flight.

The coolant air flow was achieved by connecting the suction side of asmall blower to the coolant discharge of the housing. A globe valvelocated between the condenser and the blower was used for modulatingcoolant air flow. The suction pressure at the inlet of the blower wasmeasured relative to cabin pressure using a differential pressuretransducer, and the air flow was determined from a blower curve derivedfor that specific machine. The temperature of the coolant air exitingthe condenser was measured using a K-type thermocouple, as was theincoming coolant air.

A piston pump metered liquid water to a temperature controlledmicrochannel vaporizer to produce superheated steam that was then mixedwith air flow from a mass flow controller. Additional heat was addedafter mixing using heat trace prior to entering the device. A K-typethermocouple was inserted into the header of the device to monitor theinlet temperature. A pressure transducer was used to measure the inletpressure.

The temperature of the uncondensed gas stream was measured by a K-typethermocouple in the outlet header, and the pressure was measured by atransducer connected to a tee in the outlet piping by a water filledline. Similarly, the temperature and pressure of the water outlet flowwere also monitored. Both the gas and liquid effluent streams passedthrough clear tubing located below a video camera to record breakthroughof liquid to the gas stream and gas to the liquid stream. Both streamssubsequently passed through manually operated back pressure regulatorsthat were used to regulate the pressure difference between the twostreams. This allowed for modulation of the pressure difference acrossthe pore throat.

Three-way solenoid valves were used downstream of the regulators todirect the gas and liquid flows between two collection vessels. Betweenreduced gravity experiments, both flows were directed to collectionvessels. When the g-level dropped below a preset value, normally 0.2 g,a timer counted down from 2 seconds before the solenoids were activatedto direct the flows to two sample jars. The sample jars each contained apreweighed piece of laminated sorbent material to collect water from thestream during the experiment. When the g-level exceeded 0.3 g, thesolenoids were deactivated automatically to direct the flow back to thecollection vessels. The samples were changed between experiments, storedin individual plastic bags, and weighed after the flights to obtain ameasurement of the average water flow rate in each stream during theexperiment.

In summary, phase separation efficacy was determined by visualobservation during the experiments, review of video tape recordings ofthe outlet lines, and by water balance using the samples. Heat balanceswere also calculated for both the cold and hot streams using measuredtemperatures, pressures, and flow rates.

Results and Analysis

The condenser/separator was flown on four reduced gravity flights, butthe first did not yield useful data because of operational difficulties.During the other three flights, data were collected at varying flowrates, inlet temperatures, and steam fraction of the hot feed stream.Phase separation efficacy was evaluated, and heat balance analysesresulted in assessment of heat exchanger performance, as describedbelow.

Phase Separation

The objective of phase separation is to collect and remove all of theliquid condensate from the flowing gas stream, while also precluding thebreakthrough of gas into the condensate outlet stream. Breakthrough ofgas into the condensate stream rarely occurred during the reducedgravity experiments. As long as the difference in pressure between theinlet and the liquid outlet was maintained below the breakthroughpressure of 4.4 kPa, gas was not observed in the liquid effluent. Insome experiments, gas breakthrough to the liquid stream did not occureven when the pressure difference exceeded the breakthrough pressure.Furthermore, when gas did breakthrough across the pore throat into theliquid stream, phase separation could be restored simply by modifyingoperating parameters, such as by slightly increasing the liquid outletpressure.

Liquid breakthrough to the gas outlet occurred more frequently duringthe reduced gravity experiments, although complete separation of theliquid did occur at condensation rates as high as 10 mL/min at an airflow of 5 SLPM. When the air feed rate was increase to 11 SLPM, as muchas 6 mL/min of condensate could be completely removed from the gas flowchannels.

Pore throat flow capacity has been found to be an important parameter inthe operation of microchannel phase separators (TeGrotenhuis andStenkamp, 2001). Pore throat capacity is calculated from Darcy'sequation as

$\begin{matrix}{Q_{PT} = \frac{{KA}\;\Delta\; P}{\mu\; h}} & (1)\end{matrix}$where Q_(PT) is the maximum flow of fluid having viscosity μ througharea A of a porous media having permeability K and thickness h under anapplied pressure drop of ΔP. The pore throat capacity is calculated fora given experiment using an average of the pressure difference betweenthe inlet and the liquid outlet and the pressure difference between thegas outlet and the liquid outlet.

The condensation rate is calculated by water balance. The amount ofwater vapor remaining in the gas stream assuming saturation at theoutlet temperature and pressure is subtracted from the water feed rateto arrive at a condensation rate. The ratio of the condensation rate tothe pore throat capacity is defined as the pore throat utilization. Pastwork in microchannel phase separators has found that the occurrence ofliquid breakthrough is correlated to the pore throat utilization.

Compiled liquid breakthrough results are shown in FIG. 2 for all of thereduced gravity experiments. Typically, breakthrough of condensate didnot occur when the air flow was 5 SLPM or below even as the pore throatutilization reached 25%. At the higher air flows up to 11 SLPM, liquidbreakthrough was inconsistent and occurred even at pore utilizationbelow 5%.

The amount of liquid detected in the gas stream samples was minimal atthe lower air flows of 5 and 7 SLPM; the maximum being only 3% of thewater feed. At 8 SLPM air flow, liquid measured in the gas samplesreached 13% of the liquid feed, and above 8 SLPM, liquid breakthroughapproached 40% of the water feed rate. Therefore, phase separation wasvery effective at the lower gas velocities, but became increasinglyunstable at higher velocities. Liquid entrainment was also found to besensitive to the gas outlet temperature as shown in FIG. 3. At a givenair flow rate, the likelihood of liquid breakthrough to the gas outletincreased as the temperature of the effluent gas increased.

Heat Exchange

Heat exchange effectiveness is analyzed by performing heat balances onboth the hot and cold streams. The amount of heat acquired by thecooling air stream is calculated from the temperature rise and the molarflow rate as determined from the blower curve using the pressure riseacross the blower. The amount of heat transferred from the condensingstream is the sum of the latent heat of condensation and the sensibleheat loss. The latent heat rate is calculated from the condensationrate, which is calculated as the difference between water feed rate andthe water vapor remaining in the gas stream, assuming saturated air atthe gas outlet temperature and pressure. The sensible heat loss isdetermined by cooling the condensate to the water and/or gas outlettemperatures plus cooling the outlet gas stream to the gas outlettemperature. The difference between the heat duty of the hot stream andthe duty of the cooling stream is considered the ambient heat loss.

The condenser housing contributes a large thermal mass to the system,creating long temperature transients during start-up and when operatingconditions are changed. Because the system cannot be operated duringtake-off of the aircraft, the limited duration of the flight, and themandated frequency of experiments, the condenser was operated in onlypseudo-steady-state mode. This is seen most clearly in the ambient heatloss trends. For example, the first four experiments of Flight 3 wereconducted at the same operating conditions, including the sametemperature and composition of the hot feed, but the ambient heat lossdecreased from 83% of the hot stream heat duty to 61%, while the gasoutlet temperature climbed from 20° C. to 30° C., indicating that muchof the heat was being transferred to the housing instead of to thecooling stream. At other times, the ambient heat loss would becomenegative when the water feed rate was decreased, indicating that thehousing was being cooled. The average calculated heat loss for all ofthe reduced gravity experiments was 30% of the cooling duty.

The heat transfer characteristics of the device were evaluated using thehot stream heat duty despite the long-time transients caused by thethermal mass of the housing. Average heat fluxes, calculated by dividingthe hot stream heat duty by the primary heat exchange surface area, areplotted versus mol % water in the hot feed for several feed air flows inthe condensing feed mixture as shown in FIG. 4. Average heat fluxesrange from 1 up to almost 7 W/cm².

The mean overall heat transfer coefficient, U_(m), was calculated bydividing the average heat flux by the LMTD for cross-flow heat exchangeusing inlet and outlet gas temperatures. The mean overall heat transfercoefficient is found to be a function of the water content of the feedas seen in FIG. 5. This is expected for a partial condenser because theenthalpy of the condensing stream is not a linear function oftemperature. Above the dew point, the slope is equal to the heatcapacity of the gas stream. Below the dewpoint, a latent heat ofcondensation component dominates the enthalpy change, dramaticallyincreasing the slope of enthalpy versus temperature. Furthermore, theslope is not constant below the dew point, because the condensation rateis a decreasing function of temperature.

The heat transfer productivity of the device is shown in FIG. 6. Thermalenergy density and specific power are calculated using the hot streamheat duty and the size and weight of the heat transfer componentsonly—two of the condensing elements shown in FIG. 1 and gaskets. Thevolume of the two aluminum condensing elements is 35 cm³, and the massis 52 g. In the same way as the heat flux, the heat transfer energydensity and specific power depend on the water content of the feed. Theformer ranges between 2 and 12 W/cm³ and the latter between 1200 and8000 W/kg. These values decrease if the separator is added to the sizeand weight, which has a volume of 24 cm³ and a mass of 160 g. However,the separator was fabricated out of stainless steel, and the weightcould be reduced to 50 g, if it was made of aluminum. Other lighterweight materials could also be used to reduce weight when aluminum wouldnot be appropriate, such as in corrosive applications.

Discussion

Phase separation with the microchannel partial condenser performed well,particularly at the lower gas velocities. Feed air flow in the hotstream had the greatest impact on the occurrence of liquid breakthroughto the gas outlet. Condensation occurs on one side of the gas flowchannels adjacent to the cooling channels, but is removed on theopposite side into the sorbent material and through the pore throat.This requires the condensate to migrate from the condensing side tosorbent side of the microchannels while gas is flowing down thechannels. The result is the potential for liquid to be entrained orswept by the gas toward the gas outlet before it can be effectivelyabsorbed. This potential increases with increasing gas flow.

Microchannel phase separation testing has shown a strong dependence onthe pore throat flow capacity or pore throat utilization (TeGrotenhuisand Stenkamp, 2001). This dependence was not a significant factor in theresults of these tests. Experiments were conducted with no liquidbreakthrough at pore throat utilizations as high as 25%, whilebreakthrough did occur with pore utilization below 2%. Pore throatcapacity did not appear to be a limitation in the performance of thisdevice, indicating other physical processes limited liquid flowcapacity, such as transport of condensate to the wicks and/or flow ofcondensate through the wicks to the pore throat windows.

On the other hand, the gas outlet temperature did seem to affect theoccurrence of liquid breakthrough. Liquid was more likely to be found inthe gas stream as the outlet temperature was increased. This isattributed to a higher condensation rate at the outlet end of thedevice. As the gas outlet temperature increases the temperature drivingforce increases, thereby increasing the heat flux. The higher heat fluxnear the outlet results in higher condensate production near the outlet.Furthermore, the amount of condensate produced per increment oftemperature change decreases with decreasing temperature; lesscondensate is produced going from 30° C. to 25° C. than from 40° C. to35° C. This also contributes to higher condensate production near theoutlet as the gas outlet temperature increases. Producing morecondensate near the gas outlet increases the challenge of removing thecondensate from the gas flow before it exits the gas channel.Consequently, liquid breakthrough becomes more problematic as the gasoutlet temperature increases.

The mean overall heat transfer coefficient for air-cooled partialcondensation reached 500-2000 W/m² K. These values were achieved bycreating a large extended surface area for air side heat exchange,obtaining a high fin effectiveness by using aluminum, and achieving avery small hydraulic diameter to increase the convective heat transfercoefficient on the coolant side. This resulted in heat transfer powerdensities exceeding 10 W/cm³ and specific power over 5000 W/kg. Theability to achieve these levels of hardware productivity with gas heatexchangers has numerous applications where size and weight are critical.

CONCLUSIONS

A microchannel partial condenser with integrated phase separation hasbeen successfully tested in reduced gravity aboard NASA's KC-135aircraft. Mixtures of air and steam were fed at temperatures of 70-95°C. and cooled to less than 40° C. in an air-cooled cross-flow heatexchanger. The resulting condensate was successfully separated from thegas stream in the absence of gravitational forces over a range ofoperating conditions.

Breakthrough of the condensate to the gas outlet tended to occur athigher condensing stream flow rates and was also affected by the gasoutlet temperature. The flow capacity of the separator was not asignificant factor.

The air-side heat transfer resistance was reduced through the use ofextended surfaces and by reducing the hydraulic diameter through the useof minichannels, resulting in heat transfer energy densities exceeding10 W/cm³ and specific powers over 5000 W/kg. This corresponded to heatfluxes approaching 7 W/cm².

The use of microchannels on both sides of the partial condensing heatexchanger offers a significant advantage for compact systems where sizeand weight are critical. Furthermore, the ability to integrate phaseseparation and operate independent of gravity is a significantcapability, especially for space applications.

Example 2

A second example was also a cross-flow, air-cooled microchannel partialcondenser. This device differed in that gravity was used to motivatefluid through wick structures to a pore throat structure located in theoutlet gas header. This device was comprised of a stack of heat exchangeelements, gas flow channels, and wick structures. The bottom of thestack is one of the heat exchange elements described in Example 1 and isshown at the bottom and top of the stack shown in FIG. 1. The topsurface of the heat exchange element has five (5) channels that are0.020 inch (0.05 cm) deep extending from one header to the other. Agasket was placed on top of the heat exchanger element that extendsaround the perimeter. A wick structure is placed within the gasketextending between the headers and substantially covering the fivechannels, thereby forming flow channels for the gas stream. The wickstructure consists of a layer of Delker expanded metal screen, cottoncloth material, and a second layer of Delker screen, all sewn togetherto form an integral structure.

A second heat exchange element is placed on top of the first wickstructure. This second heat exchange element is similar to the firstexcept that the 0.020 inch (0.05 cm) deep channels are formed on bothsides, and the cooling flow slots are 0.200 inch (0.51 cm) tall—twice astall as the first heat exchange element giving twice the extendedsurface area for heat transfer. Another gasket and wick structure wereplaced on the second heat exchange element, followed by a third heatexchange element with dimensions the same as the second. A third gasketand wick structure are placed over the third heat exchange element.Finally, a fourth heat exchange element is placed on top of the thirdgasket and wick. The fourth heat exchange element has the samedimensions as the first, but is inverted. The completed stack has atotal of four (4) heat exchange elements, three (3) wick structures, andsix (6) arrays of parallel condensing flow channels. The order ofcomponents is as follows: cooling channel layer: gas flow channel: wickstructure: gas flow channel: cooling channel layer: gas flow channel:wick structure: gas flow channel: cooling channel layer: gas flowchannel: wick structure gas flow channel: cooling channel layer: gasflow channel: wick structure: gas flow channel: cooling channel layer.The entire stack is placed in a housing and compressed to seal thecondensing (gas) flow channels from the coolant channels.

This device was operated in a vertical orientation, with the gas inletheader located at the top and the gas outlet header located at thebottom. A pore throat structure is inserted into the outlet gas header.Condensate that forms during cooling of the gas stream is absorbed intothe wick structures and flows downward through the wick structure andonto the pore throat structure. Various structures can be placed in thegas outlet header to allow gas to flow out of the device whilepreventing condensate entrainment. A liquid outlet is provided by a tubethat penetrates the housing and the pore throat structure. The liquidoutlet is maintained at a lower pressure than the gas outlet to removecondensate as a stream separate from the gas stream.

Measured performance of this device is shown in FIGS. 7, 8 and 9 at airfeed flow rates of 32 and 48 standard liters per minute (SLPM) in thecondensing stream. The experimental data for air flow of 32 SLPM rangedin operating conditions, with the coolant inlet temperature ranging from19° C. to 25.5° C. and the coolant superficial velocity from 715-840cm/s. The inlet temperature of the condensing stream ranged from 48° C.to 93° C., and the condensing stream superficial velocity at theentrance ranged between 900 and 1350 cm/s. The total heat transferranged between 100 and 540 Watts, the water recovery was between 64% and85%, and the overall mean heat transfer coefficient ranged from 340 to1000 Watt/m²·K. At 48 SLPM air flow in the condensing stream, thecoolant inlet temperature ranging from 19° C. to 25° C. and the coolantair superficial velocity was 730-840 cm/s. The inlet temperature of thecondensing stream ranging from 56° C. to 87° C., and the condensingstream superficial velocity at the entrance ranged between 1400 and 1700cm/s. The total heat transfer ranged between 180 and 480 Watts, thewater recovery was between 61% and 76%, and the overall mean heattransfer coefficient ranged from 530 to 800 Watt/m²·K.

In one experiment at 32 SLPM feed air flow, the coolant air was fed at20° C. at a superficial velocity of 840 cm/s. The condensing streamentered at 77° C. at a superficial velocity of 1350 cm/s, and thecondensing stream consisted of 40 mol % water. The gas stream exited at55° C. with a pressure drop of 19 inches (48 cm) of water columnrepresenting 520 Watts of heat duty and 73% water condensation. Thecoolant also exited at 55° C. with a pressure drop of 2.1 inches (5.3cm) of water column. The overall mean heat transfer coefficient wascalculated as 990 Watts/m²·K, achieving 2000 Watts/kg specific heattransfer power density and 3.1 Watts/cm² power density.

In a second experiment at 48 SLPM air flow, the coolant air was fed at16° C. at a superficial velocity of 840 cm/s. The condensing streamentered at 84° C. at a superficial velocity of 1700 cm/s, and thecondensing stream consisted of 28.5 mol % water. The gas stream exitedat 48° C. with a pressure drop of 27 inches (69 cm) of water columnrepresenting 480 Watts of heat duty and 70% water condensation. Thecoolant also exited at 48° C. The overall mean heat transfer coefficientwas calculated as 795 Watts/m²·K, achieving 1800 Watts/kg specific heattransfer power density and 2.9 Watts/cm² power density.

A theoretical model of the example device was used to predict theperformance at the operating conditions of each of the experimental datapoints shown in FIGS. 7, 8, and 9. Trendlines were obtained by a leastsquares fit of the theoretically predicted values to quadratic equationswhich are shown in FIGS. 7, 8, and 9 to illustrate that actualperformance generally exceeded predicted performance. The theoreticalmodel is a numerical integration of a local heat transfer model based onthe Colburn-Hougen (1934) method that accounts for heat and masstransfer effects in the condensing stream. In addition, the modelaccounts for the air-side heat transfer resistance, wall resistance, andpossible condensate film resistance. The eps-NTU method (Rohsenow, etal., 1998) is used for cross-flow heat exchanger with both streamsunmixed and symmetric to determine the localized heat transfereffectiveness.

NOMENCLATURE

-   -   A—pore throat area, m²    -   K—permeability, m²    -   h—pore throat thickness, m    -   Q_(PT)—pore throat flow capacity, L/s    -   ΔP—applied pressure difference across pore throat, Pa    -   μ—condensate viscosity, Poise

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1. A method of condensing water, comprising: passing a fluid mixturecomprising water vapor into a gas flow channel in a condenser, whereinthe condenser comprises a cooling channel; a gas flow channel adjacentto the cooling channel; and a liquid flow path; wherein the liquid flowpath is adjacent to the gas flow channel; wherein the cooling channel isdefined by cooling channel walls; a primary heat transfer surfacebetween the gas flow channel and the cooling channel, wherein thissurface has an area; forming a liquid in the liquid flow path; andpassing coolant through the cooling channel with a pressure drop throughthe cooling channel of no more than 4 inches (10 cm) of water column,and at least one of the following: (1) the energy density, calculatedfrom the volume of the sum of the cooling channel and the gas flowchannel, including the volume of walls defining the cooling channel, isat least 2.0 W/cm³, or (2) the specific energy, calculated from theweight of the materials defining the cooling channel and the gas flowchannel, is at least 1000 W/kg, or (3) the overall mean heat transfercoefficient is at least 500 W/cm²·K based on the primary heat transferarea between the gas flow channel and the cooling channel, or (4) atleast 70% of the water vapor in the feed stream condenses into a liquid.2. The method of claim 1 wherein at least 60% of the water vapor in thefluid mixture is condensed to a liquid.
 3. The method of claim 2 whereinthe liquid flow path comprises a wick and an open liquid flow channel.4. The method of claim 3 wherein the fluid mixture comprises an effluentfrom a fuel cell.
 5. The method of claim 2 wherein 65 to 85% of thewater vapor in the fluid mixture is condensed to a liquid.
 6. The methodof claim 2 wherein the coolant is no colder than 15° C., and wherein thefluid mixture entering the flow channel comprises 50 mol % water orless.
 7. The method of claim 1 wherein the coolant is ambient air. 8.The method of claim 7 wherein the overall mean heat transfer coefficientis at least 500 W/cm²·K based on the primary heat transfer area betweenthe gas flow channel and the cooling channel.
 9. The method of claim 7wherein at least 70% of the water vapor in the feed stream condensesinto a liquid.
 10. The method of claim 3 wherein the thickness of thewick is less than 500 μm.
 11. The method of claim 1 further comprising awicking structure within the gas flow channel that assists in removing aliquid from the fluid mixture.
 12. The method of claim 1 occurring in alaminated device comprising at least 3 gas flow channels operating inparallel.
 13. The method of claim 7 wherein net flow of coolant throughthe cooling channel is cross-flow with respect to net flow of gasthrough the gas flow channel.
 14. The method of claim 12 wherein the gasflow channels have dimensions of width and length that are at least 10times larger than the dimension of height.
 15. The method of claim 14wherein the gas flow channels are essentially planar in the region inwhich liquid is formed in the liquid flow path.